Rotor and bearing construction for rotary mechanisms



Nov. 19 1963 M1 BENTELE ETAL ROTOR AND BEARING CONSTRUCTION FOR ROTARY MECHANISMS 4 Sheets-Sheet 1 Filed May 2; 1960 In: Illlll INVENTORS MAX BENTELE CHARLES JONES ALEXANDER H.RAYE

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ATTORNEYS Nov. 19, 1963 M. BENTELE EIAL 3,111,261

ROTOR AND BEARING CONSTRUCTION FOR ROTARY MECHANISMS Filed May 2, 1960 4 Sheets-Sheet 2 a 9J- 0 mil l J L l i '5: o: W 1 j i i J m m 2 I; w v 3% o I I 22 2% L 1 m o m m q' S 3 W i i l i ll m N a l 1 I, L) I INVENTORS 1 MAX 'BENTELE CHARLES JONES ALEXANDER H.RAYE

ATTORNEY S Nov. 19, 1963 M. BENTELE ETAL 3,111,261

ROTOR AND BEARING CONSTRUCTION FOR ROTARY MECHANISMS 4 Sheets-Shei 3 Filed May 2. 1960 INVENTORS MAX BENTELE CHARLES JONES ALEXANDER H.RAYE

ATTORNEYS Nov. 19, 1963 M. BENTELE ET AL 3,111,261

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United States Patent RQTAQY MECHANlSh/id Max Bentele, Ridgewood, and .lones and Alexunder H. Raye, Paramus, NJ assignors to Qurtns- Wright Corporation, a co poration of Delaware Filed May 2, 1961'), Ser. No. 26,233 Claims. (fil. Hit-2537) The present invention relates to a rotor construction for rotary mechanisms using a lightweight metal alloy or other material which preferably has good heat conducting properties, such as a magnesium or aluminum alloy; and more particularly to means which make the use of such a rotor feasible, with consequent savings in weight and advantages in cooling efiiciency.

Although this invention is applicable to and useful in almost any type of rotary mechanism which operates at an elevated temperature; such as combustion engines, fiuid motors, fluid pumps, compressors, and the like, it is particularly useful in rotating combustion engines. To simplify and clarify the explanation of the invention the description which follows will, for the most part, be restricted to the use of the invention in a rotating combustion engine. It will be apparent from the description, however, that with slight modifications which would be obvious to a person skilled in the art, the invention is equally applicable to other types of rotary mechanisms which operate at elevated temperatures.

The present invention is particularly useful in rotating combustion engines of the type which comprise an outer body having an axis, axially-spaced end walls, and a peripheral wall interconnecting the end walls; the inner surfaces of the peripheral wall and end walls form a cavity, and an inner body or rotor is mounted within the cavity between its end walls. The inner surface of the peripheral wall is preferably parallel to the axis of the cavity and, as viewed in a plane transverse to this axis, the inner surface preferably has a multi-lobed profile which is substantially an epitrochoid.

The axis of the rotor is eccentric from and parallel to the axis of the cavity of the outer body, and the rotor has axially-spaced end faces disposed adjacent to the end walls of the outer body, and a plurality of circumferentiallyspaced apex portions. The rotor is rotatable relative to the outer body such that the apex portions substantially continuously engage the inner surface of the outer body to form a plurality of working chambers which vary in volume during engine operation, as a result of relative rotation between the rotor and the outer body.

Such engines also include an intake passage for administering a fuel-air mixtu e to the chambers, an exhaust port for the chambers, and suitable ignition means so that during engine operation the worxing chambers of the engine undergo a cycle of operation which includes no four phases of intake, compression, expansion, and exhaust. The cycle of operation is achieved as a result of the relative rotation of the inner rotor and outer body, and for this purpose both the inner rotor and outer body may rotate at different speeds, but preferably the inner rotor rotates While the outer body is stationary.

For enicient operation of the engine, its working chambers should "be sealed, and therefore an effective seal is provided between each rotor apex portion and the inner surface of the peripheral wall of the outer body, as well as between the end faces of the inner rotor and the inner surface of the end walls of the outer body.

One embodiment of the rotating combustion engine which has been successfully used in practice is an engine in which the multi-lobed inner surface of the outer body has substantially the geometric form of an epitrochoid.

ice

The shape of the inner rotor illustrated in the drawings resembles in its general configuration the shape of a theoretical rotor having the maximum size which the epitrochoidal inner surface of the outer body can accommodate. The rotor need not have this maximum size but could be cut back between its apex portions, e.g., to reduce the compression ratio of the engine. In this embodiment the shape of the rotor is polygonal or has the general configuration of a polygon having convex or concave curved arcuate sides, the apex portions of the rotor being in continuous contact with the inner wall or surface of the stator.

For purposes of illustration the following description will be related to the present preferred embodiment of the engine in which the inner surface of the outer body defines a two-lobed epitrochoid, and the rotor or inner body has three apex portions and is generally triangular in cross section but has curved or arcuate sides. It is not intended that the invention be limited, however, to the form in which the inner surface of the outer body approximates a two-lobed epitrochoid and the inner body or rotor has only three apex portions. In other embodiments of the invention the inner surface of the outer body may be substantially an epitrochoid having one less lobe than the rotor or inner body has apex portions.

In one embodiment of the invention, an end face of the rotor is provided with an internally-toothed ring gear which is concentric with the bearing bore of the rotor. An externally-toothed or pinion gear of smaller diameter than the rotor gear is secured to the outer body and adapted to fit into and mesh with the teeth of the rotor gear as the rotor makes its revolutions relative to the outer body during operation of the engine.

The bearing bore of the rotor accommodates a shaft eccentric upon which the rotor is rotatably mounted, and the shaft is coaxial with the outer body cavity. In the present invention the rotor is constructed of an aluminum alloy (or other lightweight metal alloy having good heat conducting properties) and two metal sleeves of steel or similar material are fitted into the bearing bore :of the rotor and lie between it and the eccentric when the engine is assembled. The inner of these two sleeves also carries the rotor ring gear, and this gear is registered or indexed against relative rotation with respect to the rotor by a series of radial lugs or splines which fonrn a part of the iner sleeve and ring gear, and the radial lugs of \the rotor gear are in cooperative engagement with complementary radial lugs or splines on the aluminum rotor itself adjacent to its bearing bore. Since these locating splines or lugs are radially faced, the circumferential spline toothed clearance remains constant between the steel splines of the inner sleeve and the aluminum alloy splines of the rotor in spite of the relative thermal expansions and contractions of the steel and aluminum alloy components.

This invention also provides a special notor bearing construction which accommodates the higher operating temperature and higher thermal coetficient of expansion of the aluminum alloy rotor compared to the steel shaft eccentric without excessive bearing clearance; this construction may be referred to as a semi-floating bearing. The outer steel sleeve has a sufficiently tight shrink-fit into the rotor bore that it will remain tight at operating temperatures in spite of the greater thermal expansion of the aluminum alloy rotor compared to that of the outer steel sleeve. The inner steel sleeve also has a shrink-fit into the outer steel sleeve, but it is a considerably less severe shrinlofit than that of the outer sleeve in the aluminum alloy rotor. Even at operating temperature the outer sleeve and aluminum alloy rotor continue to act as one assembly, but at an intermediate temperature before the operating temperature is reached the inner sleeve separates from the outer sleeve, attains its free dimensions,

3 and acts as a true floating bearing. This semifioating hearing design provides much less variation in effective bearing clearance between bearing and eccentric from room temperature to operating temperature than can be provided with a single steel sleeve permanently shrunkfit into an aluminum rotor.

The present invention, by insuring that the effective bearing clearance between rotor and eccentric remains small in spite of differential thermal expansion be ween these parts, provides the following positive, advanta eous, and beneficial results:

(1) A small effective bearing clearance is necessary to preserve the oil film on the bearing and to provide an operationally good bearing;

(2) Mechanical interference between the rotor and the outer body might result if the bearing clearance were to become too large.

The instant invention, in making it possible to use a lightweight metal alloy rotor mounted for rotation upon the eccentric of a steel shaft, permits the accomplishment of the following beneficial and advantageous results:

(1) Great savings in rotor weight, as compared with conventional steel or cast iron rotor constructions;

(2) The use of a lightweight metal alloy rotor permits a lightweight metal alloy outer body to be used in combination with it.

If the outer body is constructed of a lightweight metal alloy having high heat conducting properties, such as an aluminum alloy, then the rotor itself must also be constructed of such a lightweight metal alloy to pro ide a continuously small clearance between the end daces of the rotor and the end walls of the outer body. If the outer body were constructed of such a lightweight metal alloy, and the rotor were not similarly constructed, the end walls of the outer body, upon thermal expansion, would grow away from the rotor and provide too great a clearance between the rotor and the outer body to permit elfective sealing between these two components.

Also, the present invention by providing means to index or register the rotor gear against rotation relative to the rotor prevents mechanical interference between the rotor and outer body which might otherwise occur because of improper or faulty indexing of the rotor gear with respect to the rotor.

Accordingly, it is a primary object of the instant invention to provide a rotor of lightweight material, and preferably with good heat conducting properties, to minimize thermal gradients within the rotor and to provide a rotor which may nevertheless be effectively mounted for rotation upon the eccentric of a steel shaft.

It is another object of the instant invention to provide a floating rotor hearing at operating temperature to give higher load carrying capacity and to achieve self-alignment and accurate roundness of the bearing.

Another object of the present invention is to provide means for minimizing the variation of bearing clearance between shaft eccentric and rotor resulting from difierential thermal expansion as the rotary mechanism is elevated from ambient to operating temperature.

Another object of the present invention is to provide for a rotor bearing construction that at operating temperature will yield a bearing clearance with the shaft eccentri that is a function only of the bearing free diameter wi" its tolerance band and the eccentric diameter with its tolerance band.

Another object of the present invention is to provide means which at operating tem erature will virtually eliminate residual shrink-fit shape distortion or deformation of the rotor bearing caused by the shape of the rotor itself when the mechanism is cold.

Another object of the instant invention is to provide a shrink-fit double sleeve for a lightweight metal alloy rotor which has a shrink-fit consistent with the allowable stress properties of aluminum, which can be ssembled on the shaft eccentric when cold, and which will provide the desired bearing clearance between rotor and eccentric at operating temperature and keep this clearance small enough to preserve an oil film on the hearing at all times.

Another object of this invention is to provide means to properly circumferentially locate the rotor gear and to index it against rotation with respect to the rotor.

A further object of the instant invention is to provide means for registering the rotor gear against relative rotation with respect to the rotor which will maintain substantially the same desired circumferential clearance between the registering means and the rotor when the mechanism is at its operating temperature as exists when the mechanism is cold.

A still further object of the instant invention is to provide a sealing means to prevent oil leakage at the rotor end face from the operating clearance between the ou e and inner sleeves which sealing means will also ac ommodate for axial growth of the aluminum rotor relative to the steel inner sleeve as the temperature of the mechanism is elevated.

For the rotor bearing to operate properly as a hearing, it, of course, must have a certain required clearance. if the aluminum alloy rotor were mounted directly on the shaft eccentric, however, the thermal expansion of the aluminum between ambient and operating temperature might be such that the bearing clearance would increase considerably beyond acceptable operating limits.

A fundamental purpose of the present invention, thus, is to provide a means of restricting the growth of the caring clearance from ambient temperature to operating temperature to as small an increase as possible and at the same time provide a means which will register the rotor against rotation with respect to the rotor gear.

Preferabl the shaft eccentric and both the inner and outer sleeves are of steel. If the eccentric and sleeves are fairly close to the same temperature (which they normally are) and if the sleeves were not restricted, restrained, or forced to grow in any other manner than as steel itself would grow, the double sleeve and the eccentric would maintain a fairly constant clearance as the mechanism warms up to operating temperature.

It is not possible, however, to completely free the growth of the bearing bore of the double steel sleeve from the influence of t e different rate of growth of the aluminum alloy rotor. By the double sleeve arrangement, this difference in rate or" growth can be minimized, but the double steel sleeve cannot be completely isolated or divorced from influence by the aluminum alloy rotor.

During initial warm up, the inner diameter of the shrinkit double steel sleeve will grow faster than a free steel sleeve would. When, however, the inner sleeve becomes a free member, attains its free dimensions, and separates itself from the outer steel sleeve and aluminum rotor at some intermediate temperature on route between ambient temperature and operating temperature, it will grow from that temperature on as a free steel member.

When the inner sleeve starts to float (i.e., frees itself from the outer sleeve), its radial faced splines which interlock with complementary radial splines on the aluminum rotor will serve to register and properly orient the now floating inner sleeve and rotor gear relative to the rotor. The circumferential clearances between the interlocking radial splines remain constant, if uniform temperature is assumed, because the growth of both the inner sleeve and the aluminum rotor is along radial lines; the growth of both these components is proportional to the radius at the point of gr wth, and the components remain essentially concentric.

The circumferential clearance between the aluminum alloy and steel radial splines is designed to be of suflicient magnitude that it will not restrict the radial floating action of the floating hearing or inner sleeve at operating temperature. The shrink-fit between the outer sleeve and the aluminum rotor is suficiently severe to be preserved at operating temperature so that the outer sleeve and aluminum rotor always act as a single component or assembly and have no clearance between them.

Broadly described, the present invention includes means for minimizing variations in bearing clearance at all temperatures between ambient and operating temperatures which is caused by different rates of thermal expansion for the rotor and eccentric and also includes means for maintaining the relative rotational position of the rotor with respect to its floating bearing and gear at elevate temperatures.

Additional objects and advantages of the invention will be set forth in part in the description which follows, and in part will be obvious from the description, or may be learned by practice of the invention, the objects and advantages being realized and attained by means of the instrumentalities and combinations particularly pointed out in the appended claims.

The invention consists in the novel parts, constructions, arrangements, combinations, and improvements shown and described.

The accompanying drawings, which are incorporated in and constitute a part of this specification, illustrate one embodiment of the invention and, together with the description, serve to explain the principles of the invention.

Or" the drawings:

FIG. 1 is a side elevation of the mechanism with the end wall of the outer body removed to show the rotor positioned within the outer body. Portions of the rotor and outer body are shown partially in section, and running bearing clearances and clearances between splines are shown and exaggerated for clarity;

FIG. 2 is a central vertical section of the mechanism taken along the line 2-2 of FIG. 1 and in which the eccentric is shown only partially in section; running clearances are shown in an exaggerated manner for clarity;

H6. 3 is a partial central vertical section of the mechanism in which, for clarity, a number of the parts are horizontally displaced;

FIG. 4 is a diagrammatic view of the built-up rotor as an aid in describing the function of the radial splines in registering the rotor bearing and gear against relative rotation with respect to the rotor; running clearances are shown in an exaggerated manner for clarity;

FIG. 5 is a schematic diagram to help clarify the function of the semi-floating bearing in maintaining an almost constant clearance between rotor and eccentric as the mechanism warms up from ambient to operating temperature.

It is to be understood that both the foregoin general description and the following detailed description are exemplary and explanatory but are not restrictive of the invention.

Reference will now be made in detail to the present preferred embodiment of the invention, an example of which is illustrated in the accompanying drawings. As shown in FIG. 1, a generally triangular rotor having arcuate sides is eccentrically supported for rotation within an outer body 12.

Although in the illustrative embodiment shown in the drawings the outer body 12 is fixed or stationary, a practical and useful form of the invention may be constructed in which both the outer body and rotor are rotar but the eccentric is stationary; in this latter form of the invention the powershaft is driven directly by rotation of the outer body and the inner rotor rotates relative to the outer body.

A still third form of the invention is possible in which the inner body or rotor is stationary, and the outer body and eccentric are rotatable.

As shown in FIGS. 1 and 2, the rotor 1t rotates on an axis 14 which is eccentric from and parallel to the axis 16 of the curved inner surface of the outer body 12. The curved inner surface 18 of the outer body 12 is substantially an epitrochoid in geometric shape and includes two arched lobe-defining portions, or lobes. An intake port 219 is arranged to communicate with one lobe of the epitrochoidal inner surface 18, and an exhaust port 22 is arranged to communicate with the other lobe.

There are two points of least radius on the epitrochoid from its center 16. A line which connects these two points of least radius and passes through the center of the epitrochoid is designated its minor axis 24. Similarly, the epitrochoid has two points of greatest radius, and a line connecting these two points and passing through the center of the epitrochoid is designated its major axis 26.

As embodied, it is apparent that the minor axis 24 divides the epitrochoid into two halves. For convenience, the half or lobe which communicates with the exhaust port may be called the exhaust lobe and the half or lobe which communicates with the intake port may be called the intake lobe.

As embodied, the generally triangular shape of the rotor it) corresponds in its general configuration to the maximum profile or" the rotor which permits interference iree rotation of the rotor llil with respect to the outer body 12.

As shown in FIGS. 1 and 2, the outer body 12, which is stationary in this embodiment, comprises two end walls 23 an 3t and a peripheral wall 32 interconnecting these end walls. A crankshaft 34 is rotatably supported by the end walls 28 and 3b of the outer body 12 by means of conventional bearings, and the axis of the crankshaft 34 is coincident with the axis 16 of the outer body 12.

An eccentric 36 is rigidly mounted on and forms an integral part of the crankshaft 34; the axis of the eccen tric 36 is eccentric from and parallel to the crankshaft axis 16. T he rotor llfl is rotatably supported upon the eccentric 36, and the central axis of the eccentric 36 is coincident with the axis 14- of the rotor lib.

As shown in P16. 1, the rotor It) includes three apex portions 38 which carry radially movable sealing members The sealing members 4t? are in substantially continuous sliding, gas-sealing contact with the inner surface 18 of the outer body 12 as the rotor it rotates within and relative to the outer body 12.

By means of the relative rotation of the rotor 10 to the outer body 12, three variable volume working chambers 42 are formed between the outer peripheral working faces of the rotor 10 and the inner surface 18 of the outer body 12. As embodied in FIG. 1, the rotation of the rotor relative to the outer body is counterclockwise and is indicated by an arrow.

A spark plug is mounted in the peripheral wall 32 of the outer body 12, and at the appropriate time in the engine cycle, the spark plug as provides ignition for a compressed combustible mixture which upon expansion drives the rotor in the direction of the arrow. The eccentri ity (e) of the rotor axis 14 from the outer body axis 16 acts as a crank arm or moment arm to convert the energy of the expanding gases into torque on the crankshaft 34.

As the rotor it? rotates, fresh combustible gases are drawn into the working chambers 42 through the intake port Ztl. These combustible gases, or the fuel-air mixture, are then successively compressed, ignited, expanded, and finally exhausted through the exhaust port 22. All four successive phases of the engine cycle: intake, compression, expansion, and exhaust, take place within each one of the variable volume working chambers each time the rotor completes one revolution within the outer body, or for each revolution of the rotor, the engine goes through a complete cycle.

ln accordance with the invention, means are provided to minimize the variation in effective bearing clearance as the mechanism changes in temperature from ambient to operating temperature.

As embodied, and as may be most clearly seen in FIG. 3, this means comprises a pair of steel sleeves and 52. The bearing bore 43 of the aluminum rotor 10 is lined with a steel outer sleeve 5d which has a severe shrinka lit in the aluminum bearing bore The shrinlofit between the bearing bore 43 and the outer sleeve is so severe that it approaches the limit of elasticity of the aluminum rotor 1% beyond which plastic deformation of the aluminum would occur.

In practice, the outer sleeve is cooled below the ambient temperature, and the aluminum rotor 1% is raised above the ambient temperature to permit assembly of the interference-fit rotor 1d and sleeve The cold outer sleeve 59 is placed in the bearing bore 43 of the hot aluminum rotor, so that when the parts return to the ambient temperature a severe shrink-fit results. This shrink-fit is sufiicient-ly severe so that even when operating temperature of the mechanism is reached, the outer sleeve 5% keeps its shrink-fit in the bearing bore 43 and the outer sleeve 5% and aluminum rotor 1%} act as one assembly throughout the full range of temperatures of the mechanism.

Also shown in FIG. 3 is a steel inner sleeve 52. In assembly of the rotor lit), the inner steel sleeve 52 also has a shrink-fit in the rotor it or more precisely, the inner sleeve 52 has a shrink-fit in the inner diameter of the outer sleeve 5'8. The shrink-fit between the inner sleeve 52 and the outer sleeve 5 however, is considerably less severe than the shrink-fit between the outer sleeve 5t and the bearing bore 43 of the aluminum rotor ill but is suificient to ensure that at ambient temperature the inner sleeve 52, the outer sleeve 5%, and the aluminum rotor 10, will all act as one assembly.

When the mechanism begins to warm up, however, during operation, at some in ermediate temperature between ambient temperature and final operating temperature, the relatively slight shrink-fit of the inner sleeve 52 and the outer sleeve 54% is overcome by the radially outward expansion of the assembly of the outer sleeve 59 and the aluminum rotor 19. When this assembly separates itself by thermal expansion from the inner sleeve 52, the inner sleeve 52 attains its free dimensions for the intermediate temperature at which the separation occurs and continues to expand thermally as a free steel member, whereas the assembly of outer sleeve and rotor continues to expand thermally at a greater rate, which is some rate of expansion intermediate between the rates for free steel and free aluminum.

At operating temperatures beyond the intermediate temperature at which separation occurs, the inner sleeve 52 acts as a true floating bearing, since at those temperatures there will be a bearing clearance between the inner sleeve 52 and the outer sleeve 5% as well as the normal hearing clearance between the inner sleeve 52 and eccentric 36 which exists even at room temperature.

In accordance with the invention means are provided to register the rotor gear against relative rotation with respect to the rotor. An internally-toothed ring gear 54 is secured to the inner sleeve 52, and as embodied, the means for registering or indexing the rotor gear 5 3- against rotation relative to the rotor it? comprise radially faced splines or lugs 56 which are secured to the inner sleeve 52 and mesh with correspondingly shaped splines se.

The inner sleeve or ring gear 54 is in mesh or engagement with an externally-toothed or pinion gear 53 which is rigidly attached to the outer body 12. The gear ratio between the rotor gear 54 and outer body gear 58 is 3 :2, so that for every revolution of rotor in about its own axis 14, the crankshaft 34 rotates three times in the same direction.

The purpose of the gearing between the rotor and the outer body is to register or index the rotor in its position within the outer body relative to the outer body, and to relieve the load on the apex portions 38 and sealing members 41? of the rotor, since these parts would otherwise bear the load of determining the registration of the rotor and the outer body. The gearing, thus, does not function to impart torque to the crankshaft 34, which is accomplished through the eccentric 36.

As shown in H6. 1, the radial splines as of the inner sleeve 52 mate with complementary radially faced splines or lugs 6b in the aluminum rotor lil itself. The interlocking engagement between the inner sleeve splines 56 and the rotor splines as serves to register the inner sleeve 52 and rotor gear 54 against rotation with respect to the rotor at operating temperatures when the cold shrink-fit between the inner seeve 52 and outer sleeve 5%} has been overcome by elevation of temperature and the inner sleeve 52 acts as a free floating bearing.

Since both the inner sleeve splines 56 and the rotor splines 66 are radially faced and since both the inner sleeve 52 and rotor it? grow thermally along radial lines, i.e., their thermal growth is proportional to the radius at the point or" growth, the circumferential clearance between the inner sleeve splines 56 and the rotor splines 6%} remains constant regardless of temperature.

In accordance with the invention an adjustable nut 62 having a radially outward extending flange 64 is threadedly engaged with the outer diameter or" the inner sleeve 52 (see FIG. 2) on the axial end of the inner sleeve 52 0pposite the end which carries the ring gear 5 The purpose of the adjustable nut 62 with its fiange 64 and the shoulder as of the inner sleeve 52 is to seal lubricating oil from leaking out along the end face 66 of the rotor lb from the clearance between the inner sleeve 52 and the outer sleeve 56 when the engine is operating. This steel flange 64- is suiiiciently thin to permit resilient deflection to allow for axial thermal expansion or" the outer sleeve 5% and aluminum rotor 16.

Channels 45 are cut in the working faces 44 of the rotor ll) both to r duce the compression ratio or" the engine and to permit the flow of combustible gases between the intake lobe and exhaust lobe of the working chamber 42 when it is located in firing position opposite the spark plug 46, as shown in FIG. 1.

Passages for cooling and lubricating oil and cooling fluid, such as water, are appropriately arranged throughout the various parts of the mechanism.

FIG. 4 is a diagrammatic or schematic view of the built-up aluminum rotor, eccentric, and crankshaft 34. This view shows the running clearances, exaggerated for clarity, between the parts depicted and its purpose is to clarify the need for maintaining a relatively constant circumferential clearance between the inner sleeve splines 56 and rotor splines 6%.

For simplification, only four of the inner sleeve splines 56 are shown in solid line. These radially-faced inner sleeve splines 56 dovetail or mesh with radially-faced splines 61 in the rotor 10. As shown in FIG. 4, when the mechanism is at operating temperature, the inner sleeve 52 acts as a floating hearing so that there is a finite clearance between the inner sleeve 52 and the outer sleeve 5% as well as the usual bearing clearance between the inner sleeve 52 and the eccentric 36.

In an embodiment of the rotary mechanism which has been used in practice, the diametral running clearance between the inner and outer sleeves and the diametral running clearance between the inner sleeve and the eccentric are approximately equal, and each of these clearances is equal to about 0.2% of the diameter of the eccentric.

The floating action of the bearing presents a problem in achieving registraiton of the inner sleeve against rotation with respect to the rotor by means of the interlocking splines 56 and 61;: (FIG. 1), since when a load or force is applied to the rotor, as when a Working face is undergoing the expansion phase of the engine cycle, the floating bearing clearance will essentially disappear, or be taken up, on the side of the rotor adjacent to the direction from which the force is applied.

In FIG. 4, this force is represented by the arrow designated P which is shown acting radially against the working face 44. It is apparent from FIG. 4 that there will be a small amount of motion when the bearing clearances are taken up as the load or force is applied to the working face 44. This load or force F is reacted on the diameter of the eccentric 36, but the rotor must move sub stantially through the bearing clearances before the load can be efiectively reacted. In other words, there is some relative radial motion between the rotor 18, the inner sleeve 52, and the eccentric 36.

From a study of FIG. 4, it is apparent that the two radial splines designated 55:: and 551), which are in line with the applied force F, do not tend to oppose the relative movement due to the bearing clearances. is also apparent from FIG. 4, however, that the two radial splines 56c and 56d, which are normal or perpendicular to the direction in which the force F is applied, will tend to oppose any motion which would lead to an eccentric condition between the inner sleeve 52, the rotor 19, and the eccentric 36. The splines 56c and 56d, thus, tend to oppose any eccentric or radial motion which would take up the bearing clearances.

If there were no circumferential clearance between the gear splines 56 and the rotor splines 69 and the applied force F were great enough, the gear splines 56 would be sheared off or overstressed, and the load of the force F would be reacted directly on the interlocking splines 56 and it instead of on the eccentric 3%.

Accordingly, a finite clearance equal to or greater than the total radial bearing clearance is provided between the splines 56 and the corresponding rotor splines as; thus, it the total radial bearing clearance is 0.21% of the diameter of the eccentric, the clearance between the radially aligned gear faces of the splines 56 and the rotor splines 6% must be equal to or greater than 0.2% of the eccentric diameter. The provision of this constant circumferential clearance between gear splines 56 and rotor splines 6% permits the entire radial bearing clearance to be taken up or eliminated without placing any stress on the splines, and permits the load to be reacted directly on the eccentric 36.

In the actual mechanism, of course, there are considerably more splines than the four which are shown schematically in FIG. 4. These additional splines appear in intermediate positions and are shown in FIG. 4 by broken lines. As intermediate splines, they are also affected by the elimination of the radial bearing clearances when the load is reacted on the eccentric, but the clearance between these splines and their complementary splines in the rotor are affected in varying degrees, roughly sinusoidally, according to their location relative to the direction of the force F.

From FIG. 4 it is also apparent that there must be a clearance between the circumferential portions '76 and 73 of the gear splines 56 and rotor splines 6i respectively, equivalent in amount to the clearance between the radial faces 72 and 74 of the splines 56 and as, respectively. This circumferential clearance permits the floating bearing clearance to be eliminated without interference by the circumferential portions 76 and 78.

The clearance between the radial faces 72 and 74 and the equivalent clearance between circumferential portions 76 and 78 are so slight, even though they are large enough to permit the floating bearing clearance to be eliminated without interference contactor stress between the complementary splines 56 and 6%, that the registering function of the splines in keeping the inner sleeve 52 and ring gear 54 registered or indexed with respect to the rotor is for practical purposes preserved to the extent required for efficient operation of the engine.

FIG. is a sc ematic or diagrammatic representation which shows how the objects of the present invention are achieved by a built-up aluminum alloy rotor construction (comprising an aluminum alloy rotor itself, a steel outer sleeve, and a steel inner sleeve) in cooperation with an eccentric upon which the rotor is mounted for rotation. The horizontal or X axis of PH}. 5 is a plot of the temperature of the rotor in degrees Fahrenheit, and the vertical or Y axis is a plot of the relative thermal expansion or increase in size of the principal parts of the built-up rotor and the eccentric due to temperature elevation.

The t ermal expansion of each part is shown between ambient temperature and 450 F. The curves representing expansion of the parts approximate straight line equations, and they are represented in FIG. 5 as straight lines. The differential thermal expansion between parts or assemblies can readily be obtained from FIG. 5.

The lines representing the thermal expansion of the free aluminum alloy rotor and free outer sleeve are hypothetical, since, throughout the operating temperature range, the rotor and outer sleeve act as one assembly, and the line representing their combined thermal expansion is a line having a slope intermediate between the slope of the thermal expansion line for free aluminum alloy (rotor) and the slope of the thermal expansion line for free steel (sleeve).

Because of the severe shrink-fit between the steel outer sleeve and the aluminum alloy rotor, the assembled aluminum alloy rotor has a greater inner diameter than the free rotor would have, and conversely the assembled steel outer sleeve has a smaller outer dimeter than the free outer sleeve would have. At room temperature the inner steel sleeve also is not free but has a shrink-fit in the outer steel sleeve.

The slope of the equation representing thermal expansion of the complete assembly from ambient temperature (76 F.) to approximately F. is represented by the dashed line designated rotor-t-outer sleeve-i-inner sleeve assembly. At approximately 160 F. the shrink-fit of the inner sleeve within the outer sleeve is overcome and the inner sleeve from that temperature on acts as a free steel member and follows the normal thermal expansion curve for steel. The loss of the inner sleeve from the assembly at 160 F. decreases the influence of the steel on the thermal expansion curve for the assembly and, aceo-rdingh changes the slope of the curve r presenting the assembly, as can be seen in PEG. 5.

The curve of thermal expansion for the eccentric follows the normal curve for a free steel member throughout the temperature range.

The clearance at ambient temperature between the eccentric and the inner diameter of the inner sleeve is designated on FIG. 5, and the clearance at 450 F. is also designated. A comparison of these two clearances shows that their diiference is only slight, and thus one of the primary objects of the invention is achieved in minimiz ing the variation of the clearance between the inner sleeve and eccentric throughout the operating temperature range.

Ki. 5 also shows how the clearance between the inner and outer sleeves grows from a shrink-fit and no clearance at room temperature, to the beginning of a finite clearance at 160 F, and finally to a clearance between the sleeves at operating temperature which is approximately equal to the operating clearance between the inner sleeve and the eccentric. The sum of these two operating clearances is equal to the floating bearing clearance or total radial bearing clearance, and two clearances, one on each side of the inner sleeve, give the inner sleeve the possibility of eccentric radial movement and also determine the limits of that movement.

The rotor and outer sleeve assembly continuously act as one piece or one assembly throughout the range of temperatures between ambient temperature and operating temperature.

From the foregoing description, it can be appreciated that since the aluminum rotor is built-up as an assembly of a number of parts and each one of these parts has its own tolerance band, the accumulated tolerance band aifecting the clearance between the eccentric and the inner sleeve can be quite large when the mechanism is cold, or at ambient temperature.

When the mechanism is at operating temperature, hot

ever, as previously described, the inner sleeve attains its free detailed dimensions and becomes a free floating steel member, so that the only tolerance influencing the clearance between it and the eccentric is the tolerance on its inner diameter and the tolerance band of the eccentric. All of the effects of the tolerance on the outer diameter of the inner sleeve, the tolerance band of the outer sleeve, and the tolerance band of the aluminum rotor that would normally contribute to the effective tolerance of the inner diameter of the inner sleeve when the mechanism is cold, are eliminated, because the inner sleeve is free.

The operating clearance, thus, is much less aflected.

by the problem presented by tolerance bands than is the assembly or cold clearance. The operating or running clearance is, therefore, a function of only two tolerances, the tolerances on the outer diameter of the eccentric and the inner diameter of the inner sleeve.

The invention in its broader aspects is not limited to the specific mechanisms shown and described, but also includes within the scope of the accompanying claim any departures made from such mechanisms which do not depart from the principles of the invention and which do not sacrifice its chief advantages.

What is claimed is:

l. A rotor for rotary mechanisms having an axis and supported for rotation upon and relative to an eccentric portion of a member, the rotor having an inner bore; an outer sleeve having a shrink-fit into the inner bore with sutflcient tightness to maintain the shrink-fit at all temperatures between the ambient temperature and the operating temperature of the rotor; an inner sleeve bearing shrunk-fit into the inner diameter of the outer sleeve with a shrink-tit that is less tight than the shrink-fit of the outer sleeve into the inner bore; whereby the shrink-fit of the inner sleeve bearing is overcome at an intermediate temperature between the ambient temperature and the operating temperature of the rotor so that the inner sleeve bearing acts as a floating bearing between the eccentric portion and outer sleeve above the intermediate temperature.

2. The invention as defined in claim 1, in which the rotor is an aluminum alloy and the outer sleeve and the inner sleeve hearing are steel.

3. The invention as defined in claim 1, that also in- Ycludes a nut aflixed to one axial end of the inner sleeve bearing, the nut having a flange engaged in sliding and sealing engagement with one axial end of the outer sleeve, whereby the nut and flange seal lubricating fluid in the running bearing clearance between the inner sleeve bearing and the outer sleeve at one of their axial ends.

4. The invention as defined in claim 1, that also includes an internally-toothed gear rotatable with the rotor and secured to the inner sleeve bearing; the means for securing the internally-toothed gear to the inner sleeve bearing comprising external splines on the internallytoothed gear and internal splines on the rotor in interlocking engagement with the external splines with clearance both radially and circumferentially between the external and internal splines; the external and internal splines registering the rotor and internally-toothed gear against relative rotation regardless of differential thermal expansion between the rotor and internally-toothed gear as the rotor varies in temperature between the ambient temperature and its operating temperature.

5. The invention as defined in claim 4, in which the faces of the external and internal splines are radial relative to the axis of the rotor.

6. A rotor for rotary mechanisms having an axis and suppor ed for rotation upon and relative to an eccentric portion or" a member; bearing means between the rotor and eccentric portion for dividing increases in bearing clearance between the rotor and eccentric portion due to differential thermal dimensional changes of the rotor relative to the eccentric portion; the bearing means including a sleeve having a shrink-fit in the inner perimeter of the rotor sufficiently tight to remain in tight contact with the rotor at maximum operating temperatures; and an inner b adjacent to the eccentric portion and having a sk-fit within the sleeve that is less tight than the fit of me sleeve in the rotor; the shrink-fit of the inner bearing in the sleeve being overcome at an intermediate temperature between the ambient temperature and the operating temperature of the rotor; whereby the inner bearing acts as a floating bearing between the eccentric portion and sleeve above the intermediate temperature and divides the total clearance between the hearing and sleeve into two clearances: (l) a clearance between the eccentric portion and the inner bearing, and (2) a clearance between the inner bearing and the sleeve so that the clearance between the eccentric portion and the in er bearing is kept sufllciently small to insure that a film of lubricating fluid may be maintained at all times between the eccentric portion and the inner bearing.

7. The invention as defined in claim 6, in which the rotor is aluminum and the sleeve and inner hearing are steel.

8. The invention as defined in claim 6, that also includes a nut aflixed to one radial end of the inner bearing, the nut having a flange engaged in sliding and sealing engagement with one axial end of the sleeve, whereby the nut and flange seal lubricating fluid in the running "caring clearance between the inner bearing and the sleeve at one of their axial ends.

9. The invention as defined in claim 6, that also includes an internally-toothed gear rotatable with the rotor and secured to the inner bearing; the means for securing the internally-toothed gear to the inner bearing comprising external splines on the internally-toothed gear and internal splines on the rotor in interlocking engagement with the external splines with clearance both radially and circumterentially between the external and internal splines; the external and internal splines registering the rotor and internally-toothed gear against relative rotation regardless of differential thermal expansion between the rotor and the internally-toothed gear as the rotor varies in temperature between the ambient temperature and its operating temperatures.

10. The invention as defined in claim 9, in which the faces of the external and internal splines are radial relative to the axis of the rotor.

References Cited in the file of this patent UNITED STATES PATENTS 

6. A ROTOR FOR ROTARY MECHANISMS HAVING AN AXIS AND SUPPORTED FOR ROTATION UPON AND RELATIVE TO AN ECCENTRIC PORTION OF A MEMBER; BEARING MEANS BETWEEN THE ROTOR AND ECCENTRIC PORTION FOR DIVIDING INCREASES IN BEARING CLEARANCE BETWEEN THE ROTOR AND ECCENTRIC PORTION DUE TO DIFFERENTIAL THERMAL DIMENSIONAL CHANGES OF THE ROTOR RELATIVE TO THE ECCENTRIC PORTION; THE BEARING MEANS INCLUDING A SLEEVE HAVING A SHRINK-FIT IN THE INNER PERIMETER OF THE ROTOR SUFFICIENTLY TIGHT TO REMAIN IN TIGHT CONTACT WITH THE ROTOR AT MAXIMUM OPERATING TEMPERATURES; AND AN INNER BEARING ADJACENT TO THE ECCENTRIC PORTION AND HAVING A SHRINKFIT WITHIN THE SLEEVE THAT IS LESS TIGHT THAN THE FIT OF THE SLEEVE IN THE ROTOR; THE SHRINK-FIT OF THE INNER BEARING IN THE SLEEVE BEING OVERCOME AT AN INTERMEDIATE TEMPERATURE BETWEEN THE AMBIENT TEMPERATURE AND THE OPERATING TEMPERATURE OF THE ROTOR; WHEREBVY THE INNER BEARING ACTS AS A FLOATING BEARING BETWEEN THE ECCENTRIC PORTION AND SLEEVE ABOVE THE INTERMEDIATE TEMPERATURE AND DIVIDES THE TOTAL CLEARANCE BETWEEN THE BEARING AND SLEEVE INTO TWO CLEARANCES: (1) A CLEARANCE BETWEEN THE ECCENTRIC PORTION AND THE INNER BEARING, AND (2) CLEARANCE BETWEEN THE INNER BEARING AND THE SLEEVE SO THAT THE CLEARANCE BETWEEN THE ECCENTRIC PORTION AND THE INNER BEARING IS KEPT SUFFICEINTLY SMALL TO INSURE THAT A FILM OF LUBRICATING FLUID MAY BE MAINTAINED AT ALL TIMES BETWEEN THE ECCENTRIC PORTION AND THE INNER BEARING. 